Ship&#39;s hull vibration damper

ABSTRACT

There is described an apparatus for inclusion in a ship comprising a mass (10), that is vibrationally coupled to the ship&#39;s hull (12) such that movement of the mass is dampened. The mass (10) is so sized and located as to, when in use it reduces the fundamental and/or harmonics of a transverse two node mode of wave-induced vibration in the ship&#39;s hull (12). There is also described an apparatus for inclusion in a ship comprising a mass (10) made up in part of the tackle or fitments of the ship, which also are vibrationally coupled to the ship&#39;s hull (12) by a damping such that movement of the mass (10) is dampened so as to thereby, when in use, reduce vibrations in the ship&#39;s hull (12). In a preferred embodiment the mass (10) is at least in part a chain locker and that portion of the chain stored therein. By reducing vibrations in the ship&#39;s hull (12) significant stresses may be relieved and the performance of the ship enhanced. A ship incorporating the apparatus and a method of reducing vibrations in a ship&#39;s hull are also described.

BACKGROUND OF THE INVENTION

The present invention relates to an apparatus for inclusion in a ship toreduce vibrations in the ship's hull and to a ship incorporating such anapparatus. The present invention additionally relates to a method ofreducing vibrations in a ship's hull.

It is widely accepted that hull vibrations are capable of giving rise tosignificant stresses within the structure of a ship which may, in time,lead to structural fatigue and ultimately structural failure. Theelmination or reduction of these vibrations is therefore of considerableinterest to ship designers.

Hull vibrations span a very large frequency range from the movement ofthe ship in the seaway through the low frequency hull girder vibrationsand the various forms of engine vibration to the higher frequency localvibrations. The means used to eliminate or reduce these vibrations aresimilarly diverse. The movement of a ship in the seaway, for example, isgenerally counteracted by the use of stabilisers of which theDenny-Brown is probably the best known design offered commercially. Thedesign relies on the use of hydrofoils on either side of the ship whichprovide moments that compensate for the action of the waves and therebyreduce the rolling of the ship.

Hull girder vibrations on the other hand, have until recently beenthought to be caused exclusively by excitations from the main engines,propellors and other machinery contained within the ship. In particularit has been thought that these vibrations were predominantly excited bythe primary and secondary out of balance of the main engine. Balancedmain engines have therefore often been fitted to ships to minimise thesehull girder vibrations, commonly in combination with a Nishishibabalancer which is used to balance the second order out of balance of themain engine.

More recently however, a better understanding of the importance of waveexcitation has been developed and it is now recognised that lower modesof hull girder vibration may also be excited by the action of waves. Ina recent experiment conducted during a ship's sea trials a wave-excited,two node mode of vibration was detected at the after end of the ship'shull having an amplitude of 1.78 mm and a frequency of 1.48 Hz, and thison a day when the weather was good and the sea relatively calm. It wascalculated that the detected vibration gave rise to a nominal vibrationstress in the deck of 2.5N/mm² while further measurments suggested thatthe speed of the ship was adversely affected.

It is now thought that the hull girder vibrations may also be excited byprocesses known as "bottom slamming" and "bow flare slamming" in which,as waves break over the bow of the ship, the buoyancy of the bow sectionis alternately decreased and increased setting up a transverse standingwave throughout the length of the hull. The amplitude of the vibrationis a maximum at the bow where the excitation occurs but can give rise tosignificant stresses throughout the ship, particularly at deck levelboth because of the presence of hatches and also because the doublebottomed nature of a typical hull provides a much stronger structure.For example the two node mode of vibration in which the wavelength isequal to the length of the ship and both bow and stern are antinodes isthought to give rise to stresses at least as large as those calculatedon the basis of the rigid body assumption.

The frequency of hull girder vibrations are primarily determined by thestructural stiffness and mass of the hull. Because mass plays a part indetermining the frequency of the vibrations it will of course mean thatthe hull of a ship when loaded will have a different natural frequencythan when unloaded. Typically for longer ships however, the frequncy ofvibration of the two node mode is between 0.6 Hz and 1.0 Hz whilst forsmaller ships the frequency of vibration may be increased to nearer 2.0Hz. Since under typical sea conditions waves are present from a broadfrequency spectrum, it is to be expected that there will always be somewaves present capable of stimulating this two node mode of vibration. Asship designers can have no control over the excitation of thesevibrations it is considered that the best way of minimising theresulting stresses is to reduce the amplitude of the vibrations by meansof damping.

At higher frequencies, modes of hull girder vibration become difficultto distinguish from local vibrations. In this frequency range the mostimportant excitations are usually harmonics of the blade frequency, thatis the product of the number of blades provided on the propeller and itsrotational speed, and orders corresponding to the number of cylinders inthe main engine. These vibrations are relatively difficult to predictwith accuracy but as far the propeller is concerned, the vibrations maybe limited by using relatively large propeller clearances. This canhowever result in some loss of propulsive efficiency. As for the engineinduced vibrations, axial vibration dampers and various forms of enginestays are just some of the controling means available to the shipdesigner. The literature available in the art would suggest thatvibrations in this frequency range seldom cause major problems butfurther damping could clearly enhance ship development.

SUMMARY OF THE INVENTION

According to a first aspect of the present invention there is providedan apparatus for inclusion in a ship comprising a mass, means forvibrationally coupling the mass to the ship's hull and means for dampingthe movement of the mass, the mass being so sized and located as to,when in use, reduce the fundamental and/or harmonics of a transverse twonode mode of wave-induced vibration in the ship's hull.

Advantageously the mass is in part comprised of the tackle, fitments orcargo of the ship.

According to a second aspect of the present invention there is providedan apparatus for inclusion in a ship comprising a mass made up in partof the tackle or fitments of the ship, means for vibrationally couplingthe mass to the ship's hull and means for damping the movement of themass so as to, when in use, thereby reduce vibrations in the ship'shull.

Advantageously the mass comprises between approximately 0.2% andapproximately 0.5% of the unloaded tonnage of the ship. In a preferredembodiment the mass is in part comprised of a chain locker and thatportion of the chain stored therein.

Advantageously the means for vibrationally coupling the mass to theship's hull comprises a resilient support and in one embodimentcomprises at least one spring. In another embodiment the means forvibrationally coupling the mass to the ship's hull comprises a hydragassuspension system. Preferably the hydragas suspension system comprisesat least two cylinders disposed in opposed relationship.

Advantageously the mass is coupled to the ship's hull in such a way asto vibrate at a frequency within 10% of the resonant frequency. In apreferred embodiment the mass is coupled to the ship's hull in such away as to vibrate when the ship is unloaded at a frequency less than theresonant frequency of a mode of vibration that is to be damped. Thefrequency of vibration of the mass may be varied by altering themagnitude of the mass and/or by altering the stiffness with which themass is vibrationally coupled to the ship's hull.

Advantageously the means for damping the movement of the mass comprisesat least one shock absorber and in a preferred embodiment comprises atleast one piston damper.

According to a third aspect of the present invention there is provided aship incorporating an apparatus comprising a mass, means forvibrationally coupling the mass to the ship's hull and means for dampingthe movement of the mass, the mass being so sized and located as to,when in use, reduce the fundamental and/or harmonics of a transverse twonode mode of wave-induced vibration in the ship's hull.

Advantageously the mass is in part comprised of the tackle, fitments orcargo of the ship.

According to a fourth aspect of the present invention there is provideda ship incorporating an apparatus comprising a mass made up in part ofthe tackle or fitments of the ship, means for vibrationally coupling themass to the ship's hull and means for damping the movement of the massso as to thereby reduce vibrations in the ship's hull.

Advantageously the apparatus is located at a point within one eighth Ofa wavelength of an antinode of a mode of vibration that is to be dampedsuch as for example, in the case of wave-induced vibrations, within thebow region of the ship.

According to a fifth aspect of the present invention there is provided amethod of reducing the fundamental and/or harmonics of a transverse twonode mode of wave-induced vibration in a ship's hull comprising thesteps of providing a mass of sufficient size, vibrationally coupling themass to the ship's hull at an appropriate location and damping theresulting movement of the mass.

According to a sixth aspect of the present invention there is provided amethod of reducing vibrations in a ship's hull comprising the steps ofidentifying a mass made up in part of the tackle or fitments of theship, vibrationally coupling the mass to the ship's hull and damping theresulting movement of the mass.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic side view of a ship which has undergone sea trialsrelating to wave induced vibration;

FIG. 2 is a schematic plan view of the ship of FIG. 1;

FIG. 3 is a series of graphs showing the way in which the vibrationalspectrum at a particular point on the ship varies with increasing enginespeed;

FIG. 4 is a schematic illustration of a mathematical model of avibrating ship;

FIG. 5 is a schematic illustration of a simplified version of themathematical model of FIG. 4;

FIG. 6 is a graph illustrating the variation of vibrational amplitudewith frequency based on the mathematical model of FIG. 4;

FIG. 7 is a schematic view of an apparatus in accordance with the firstaspect of the present invention;

FIG. 8 is a side view of the bow portion of a ship showing the usualposition in which chain lockers are located;

FIG. 9 is a plan view of the bow portion of a ship showing the usualposition in which chain lockers are located;

FIG. 10 is a schematic illustration of a further mathematical model of avibrating ship;

FIG. 11 is a graph illustrating the variation of vibrational amplitudewith frequency based on the mathematical model of FIG. 10; and

FIG. 12 is a schematic view of a hydragas suspension system in place ofthe springs of the apparatus of FIG. 7.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Advantageously the mass is coupled to the ship's hull in such a way asto vibrate at a frequency within 10% of the resonant frequency of a modeof vibration that is to be damped. In a preferred embodiment the mass iscoupled to the ship's hull in such a way as to vibrate when the ship isunloaded at a frequency less than the resonant frequency of a mode ofvibration that is to be damped. The frequency of vibration of the massmay be varied by altering the magnitude of the mass and/or by alteringthe stiffness with which the mass is vibrationally coupled to the ship'shull.

The first stage in designing a vibration damper is to establish amathematical model of the vibration. Predictions from the model may thenbe made and, if actual measurements are available, adjustmentsincorporated so that the predicted and measured values are in agreement.A damper can then be simulated as an addition to the model and theeffect of the damper accurately predicted.

In order to obtain actual measurements with which to test a futuremathematical model, hull girder vibrations are usually measured at about12 positions throughout the length of the ship and these positions arechosen to obtain the best possible picture of the different modes ofvibration. FIGS. 1 and 2, represent respectively a side view and a planview of the ship previously referred to as having undergone sea trialsfrom which it can be seen that in this particular case, threetransducers were placed at the forward end of the ship, three at theafter end, four at the bridge top and one on the stern tube forwardgland. In addition a revolutionary marker was used to monitor thepropeller revolutions and provide a time signal.

The signals from the transducers were captured on a high quality taperecorder having the required number of channels and a frequency responsesuitable for the vibrations in question. Afterwards the signals could beplayed back in the same way as they were captured taking care to ensuretheir correct callibration.

Measurements were carried out with the ship travelling at differentspeeds and since in each case measurements were taken from 12 positions,a large volume of date was stored on the tape recorder. A properanalysis of data collected in this way can be a formidable task in thatit requires an understanding of the vibration characteristics of a shipas well as the use of the best available instruments such as FastFourier Transform analysers. FFT analysers are used to extract thesignificant vibrational frequencies from the records stored on the taperecorder and these are then further analysed using a computer. A typicalresult is shown in FIG. 3 in which the two node mode of vibration isclearly visible. The engine and propeller excited vibrations are alsoclearly distinguishable as for these vibrations the frequency isproportional to engine speed.

Turning now to the establishment of a mathematical model, the two nodemode of vibration of the ship referred to above as having undergone seatrials will be represented to a first approximation by the six masssystem shown in FIG. 4.

When considering the vibrational chacteristics of a ship it is necessaryto take into account not only the action of the surrounding water on theship but also the action of the ship on the surrounding water. This isdone by reference to the virtual mass of the ship which is the sum ofthe displacement of the ship, in the present case 32,672 tonnes, and anadded mass. The British Ship Research Association have established amethod of determining the value of the added mass in particularinstances and in the present case the virtual mass of the ship wascalculated to be 90,044 tonnes. The sum of the six masses in therepresentative system was therefore made equal to the vitual mass of theship and the stiffness of the system arranged so as to yield themeasured natural frequency of 1.04 Hz. The Holzer table shown belowgives the calculated modal shape of the system. This is to be comparedwith the measured modal shape shown as a dashed line in FIG. 1 and itwill be seen that the predicted and measured results are in agreement.

                                      TABLE 1                                     __________________________________________________________________________    HOLZER FREQUENCY TABLE                                                        II-Node: 62.5 VPM, ω 6.545 1/sec, ω.sup.2 = 42.837                1/sec.sup.2.                                                                     Mass                                                                              Mω.sup.2                                                                       Δ.sub.n                                                                     Mω.sup.2 Δ.sub.n                                                       ΣMω.sup.2 Δ.sub.n                                                K.sub.n                                                                            Δ.sub.n -Δ.sub.n-1               No.                                                                              tonne                                                                             tonne/sec.sup.2                                                                      mm  t.mm/s.sup.2                                                                       t.mm/s.sup.2                                                                       N/mm mm                                           __________________________________________________________________________    1  16371                                                                             701285 1.000                                                                             701285                                                                             701285                                                                             586000                                                                             1.197                                        2  "   "      -0.197                                                                            -138153                                                                            563132                                                                             "    0.961                                        3  "   "      -1.158                                                                            -812088                                                                            -248956                                                                            "    -0.425                                       4  "   "      -0.733                                                                            -514042                                                                            -762998                                                                            "    -1.302                                       5  "   "      0.569                                                                             399031                                                                             -363967                                                                            779380                                                                             -0.467                                       6   8186                                                                             350663 1.036                                                                             363287                                                                             -680                                                   __________________________________________________________________________

It is common practice to represent the damping of a vibration system asa dynamic amplification factor which is calculated by modelling thesystem as a single mass system as shown in FIG. 5. This approach hasbeen adopted in the past by the British Ship Research Association andusing their results it is possible to estimate that the dynamicamplification factor in the present case would have a value between 160and 320.

For the purpose of designing a vibration damper it is convenient torepresent the damping as a damping force proportional to the velocity ofthe vibration, ie. to assume that the damping is viscous. For the singlemass system shown in FIG. 5 the damping force may be represented by adashpot in parallel with the spring. The relationship between thedamping constant and the amplification factor at resonance is given by:

    C.sub.s =m·w.sub.O /k.sub.m =k/w.sub.O ·k.sub.m

where

C_(s) =damping constant,

m=mass of the system,

k=stiffness of the spring,

k_(m) =dynamic amplification factor at resonance,

w_(O) =natural frequency.

Substituting the limiting values of the amplification factor obtainedfrom the British Ship Research Association, the following values areobtained for the damping constant C_(s) :

k_(m) =160, C_(s) =3684 tonne/sec

k_(m) =320, C_(s) =1842 tonne/sec

The next stage in the preparation of data for the six mass system is tocalculate the damping constants from the damping constant obtained forthe single mass system. For this purpose it will be assumed that thevibrational energy absorbed by the single mass system is the same as thevibrational energy absorbed by the six mass system. Thus:

    C[(Δ.sub.1 -Δ.sub.2).sup.2 +(Δ.sub.2 -Δ.sub.3).sup.2 +(Δ.sub.3 -Δ.sub.4).sup.2 +(Δ.sub.4 -Δ.sub.5).sup.2 +1.33 (Δ.sub.5 -Δ.sub.6).sup.2 ]=C.sub.s

Using the data from Table 1:

    C[1.197.sup.2 0.961.sup.2 +0.425.sup.2 +1.302.sup.2 +1.33×0.467.sup.2 ]=3684

C=814 tonne/sec

Similarly, when k_(m) =320, C=407 tonne/sec.

The wave excitation can be obtained with sufficient accuracy as follows.It will be seen from the measured modal shape in FIG. 1 that for anamplitude of 1.78 mm at the transducer the amplitude at the centre of M₁will be 1.3 mm. Assuming that the wave excitation occurs at M₆ and thatit can be represented by a single force varying sinusoidally, themagnitude of this force can be calculated since:

    1.3=FΔ.sub.6 /w (4.523C)=1.036F/ (6.545×3684)

Therefore: F=30,256N

Similarly, when K_(m) =320, F=15,128N

The projected area of the bulbous bow in the horizontal plane isapproximately 30 m² The above force therefore corresponds to a pressurevariation of:

30,256/30=1009N/m² =approximately 0.01 bar or 0.1 meter of water

It seems reasonable that the waves could present this variation in headof water to the projected area of the bulbous bow in the sea conditionsreported.

This completes the treatment of the undamped vibrational model of theship shown in FIG. 1. A forced-damped calculation was performed on thesystem and the results are shown in FIG. 6. For the purpose of thiscalculation it was assumed that the applied force remained constantthroughout the frequency range and it will be seen that a maximumamplitude of 1.3 mm is reached at a frequency of 1.04 Hz which confirmsthe figures used in the vibrational model. It should also be noted theflanks are higher for higher values of the damping constant which in thepresent case correspond to higher values of applied force.

An embodiment of the present invention will now be described by way ofexample with reference to the accompanying drawings in which:

FIG. 7 is a schemtic view of an apparatus in accordance with the firstaspect of the present invention; and

FIGS. 8 and 9 are respectively a side view and a plan view of the bowportion of a ship and show the usual position of chain lockers on boarda ship.

The apparatus shown in FIG. 7 comprises a mass 10 coupled to a ship'shull 12 by means of a resilient support system 14.

The mass 10 preferably comprises between approximately 0.2% andapproximately 0.5% of the ship's unloaded tonnage. Since the inclusionin a ship of an additional mass of this order would have a detrimentaleffect on the ship's overall performance, the mass 10 is preferablycomprised of items already on board the ship and serving other purposessuch as the ship's tackle, fitments or cargo. It is to be noted howeverthat if the mass 10 is in part comprised of the cargo of the ship, thenwhen the ship is unloaded the mass 10 will be reduced in magnatude andnecessitate the use of one or more ballast tanks to compensate for thisfact.

In the embodiment shown in FIG. 7, the mass is comprised of a chainlocker and that portion of the chain stored in the locker. Typically aship may possess two anchor chains disposed on opposite sides of theship close to the bow with each anchor chain having a mass of typically44.5 tonnes. Allowing for the mass of the chain lockers therefore, thecombined mass of the two chain lockers and the chains they contain maytypically be of the order of 90 tonnes.

The resilient support system 14 used to couple the mass 10 to the ship'shull 12 may typically comprise a number of springs. 16 disposed betweenfirst and second projections 18 and 20 connected respectively to theship's hull 12 and the peripheral surface of the mass 10. The springs 16may be of any suitable design and in FIG. 7 are shown as being similarto the suspension springs used in railway carriages or heavy dutylorries.

The stiffness of the springs 16 and the magnitude of the mass 10together determine the natural frequency of the apparatus. This naturalfrequency is arranged so as to be within 20%, or more preferably 10% ,of the resonant frequency of the wave-induced mode of vibration in theship's hull 12 that is to be damped. In order to accomplish this tuningof the apparatus the magnitude of the mass 10 may be variable. Thus ifthe mass 10 were to in part comprise a chain locker and that portion ofthe chain stored therein, as shown in FIG. 7, different lengths of chainmay be stored in the locker to facilitate the tuning of the apparatus. Aweight gauge (not shown) may be provided to measure the magnitude of themass 10 thereby aiding the tuning process. Unfortunately, to change thenatural frequency of the apparatus by 10% would require a change of 20%in the magnitude of the mass 10 making such a method of tuningcumbersome and inefficient.

In a preferred arrangement therefore, the apparatus is tuned byselecting the springs 16 to be an appropriate stiffness. This ispossible because in each of the different modes of wave-inducedvibration the resonant frequency of the ship's hull 12 is primarilydetermed by its structural stiffness and is therefore a constant for anygiven ship. Thus at the time the apparatus is installed in the ship themode of wave-induced vibration to be damped may be selected and thesprings 16 chosen accordingly.

Having said that however, the mass of the ship does have a slight effecton the resonant frequency of the ship's hull 12 so the springs 16 arepreferably selected so as to have a stiffness that will result in thenatural frequency of the apparatus being less than the resonantfrequency of the mode of vibration of the ship's hull 12 that is to bedamped when the ship is unloaded. Thus when the ship is loaded theapparatus will vibrate with a natural frequency closer to the resonantfrequency of the mode of vibration of the ship's hull 12 that is to bedamped.

In practice the apparatus is deliberately tuned so as to have a naturalfrequency slightly different from that of the resonant frequency of themode of vibration of the ship's hull 12 that is to be damped in order toprevent the mass 10 from oscillating with too large an amplitude. Thisdifference in frequency obviously has an adverse effect on theperformance of the apparatus in reducing the wave-induced vibrations inthe ship's hull 12 but does provide a degree of protection for thesprings 16.

When correctly tuned, the mass 10 will oscillate driven by thewave-induced vibrations in the ship's hull 12 although the efficiencywith which the mass 10 is driven is dependant upon its location relativeto the ship. Clearly the mass 10 will oscillate with a larger amplitudewhen it is placed close to an antinode of the driving vibration and willnot oscillate at all if placed at a node. In general the apparatus willoperate efficiently if the mass 10 is disposed within one eighth of awavelength of an antinode of the mode of vibration of the ship's hull 12that is to be damped. Since different modes of wave-induced vibrationsin the hull 12 all have an antinode at the bow, this provides a furtherargument for the mass 10 to comprise at least in part a chain locker andthat portion of the chain stored therein since as FIGS. 8 and 9 show,chain lockers are commonly located close to the bow of the ship.

As shown in FIG. 7 the mass 10 is constrained to vibrate reciprocally bymeans of guides 24 disposed on opposite sides of the mass 10 and securedto the ship's hull 12. The oscillatory/motion of mass 10 will procludeliving space or sensitive machinery from being situated on the mass 10.Likewise this movement will make it difficult to provide pipe and/orelectrical connections across the springs 16 to the mass 10. Thishighlights yet a further reason why the mass 10 should at least in partcomprise a chain locker and that portion of the chain stored thereinsince such lockers tend not to have any sensitive apparatus mounted onthem and the chain is typically the only mechanical connection betweenthe locker and the rest of the ship. All that is necessary is for thelength of the chain between the locker and its associated winch 26 to beable to accommodate the oscillations of the locker.

In order to damp the wave-induced vibrations in the ship's hull 12,means 28 are provided to damp the oscillations of the mass 10. Thedamping means 28 may typically comprise one or more shock absorbersdisposed between third and fourth projections 30, 32 connectedrespctively to the ship's hull 12, and the peripheral surface of themass 10. The shock absorbers may be of any suitable design and are shownin FIG. 7 to comprise a piston damper similar to that found in trains orheavy duty lorries.

Returning to the mathematical model discussed previously, a shipincorporating an apparatus in accordance with the first aspect of thepresent invention may be represented to a first approximation by thesystem shown in FIG. 10.

As has been seen the two chain lockers are situated almost on the centreline of M₆ and have a combined mass of approximately 90 tonnes. Hencetreating the mass of M₆ separately from that of the two chain lockers:

M₆ =8096 tonne and M₇ =90 tonne

C₆ =3880N/mm

Assuming the apparatus to be tuned to the resonant frequency of the twonode mode of wave-induced vibrations in the hull 12, a series offorced-damped calculations may be performed to obtain a value for theoptimum damping constant at which the amplitude of vibration at M₁ is aminimum. The results of these calculations are shown in FIG. 11 fromwhich it can be seen that the maximum amplitude of vibration at M₁ hasbeen reduced to 0.195 mm assuming the original value of K_(m) to havebeen 320. If however, the original value of k_(m) was 160, the maximumamplitude of vibration at M₁ would have been reduced to 0.345 mm. Inboth cases the apparatus described provides a significant reduction inthe amplitude of the two node mode of vibration. The value of thedamping constant of the apparatus described was 30 tonne/sec.

The maximum movement across the apparatus under the above conditions isgiven by:

    For k.sub.m =160, Δ.sub.7 -Δ.sub.6 =6.127 mm

    For k.sub.m =320, Δ.sub.7 -Δ.sub.6 =3.478 mm

It will be remembered that this vibration gave rise to an nominal hullgirder stress of 2.5N/mm², and this in relatively good conditions. Ithas been predicted that "bottom slamming" and "bow flare slamming" couldcause stresses of the order of 300N/mm² in bad weather which wouldcorrespond to a maximum movement of:

For k_(m) =160, maximum movement=6.127×300/2.5 =735 mm

For k_(m) =320, maximum movement=3.478×300/2.5 =417 mm

In order to reduce this maximum movement a larger mass 10 or a greaterthan optimum damping might be used. Alternatively the apparatus might bedeliberately de-tuned.

In another arrangement however this maximum movement might beaccommodated by replacing the springs 16 of the embodiment shown in FIG.7 with a hydragas suspension system. One such system is shownschematically in FIG. 12 to comprise two piston rods 34 which are eachmounted for reciprocating motion within respective oil-filled cylinders36. These cylinders 36 are arranged so as to be in communication witheach other by means of an interconnecting pipe network 38 and incommunication with a further cylinder 40 by way of a valve 42. Thisfurther cylinder 40 is in turn arranged to be partly filled with oil andpartly filled with nitrogen gas.

In use the weight of the mass 10 is supported by the two piston rods 34which in turn are supported by the pressure within the oil/nitrogencylinder 40. The flexibility, and therefore stiffness, of the hydragassystem derives from the compressability of the nitrogen gas containedwithin the oil/nitrogen cylinder 40. Thus it will be apparent that byvarying the volume of the nitrogen gas contained within the cylinder 40the stiffness of the system may be adjusted to any desired value withincertain limits imposed by the dimensions of the system as a whole. Thischange in volume of the nitrogen gas may be achieved by the provision ofa piston within the oil/nitrogen cylinder 40 or by connecting theoil/nitrogen cylinder 40 to a reservoir of nitrogen gas via a secondvalve and/or pump.

Such a hydragas system could therefore not only be designed toaccommodate the maximum movement that would be likely to be encounteredbut could also provide a particularly simple way of adjusting thestiffness with which the mass 10 is coupled to the ship's hull 12 andthus the natural frequency of the apparatus as a whole.

A further advantage of the proposed hydragas system is that it is asimple matter to provide such a system with the necessary cooling tocounter the heat generated when the mass 10 is vibrating with aparticularly large amplitude. The same however is not true in the caseof a rubber suspension system or one comprising one or more mechanicalsprings.

It will be apparent to those skilled in the art that whilst a hydragassuspension system has been shown comprising two interconnectedoil-filled cyclinders 36, the number of oil-filled cylinders need not belimited in this way. Indeed any desired number of cylinders may be used.It will also be apparent that the oil-filled cylinders 36 comprisedwithin a hydragas suspension system need not necessarily be arranged inthe same orientation. Thus some of the oil-filled cylinders 36 may bearranged opposite others within the system such that as the mass 10 isdisplaced in one direction some of the piston rods 34 associated withthe cylinders move upwardly while others move downwardly. In this way itis possible to provide a suspension system having a greater range ofpossible stiffness and stroke.

It will also be apparent to those skilled in the art that a hydragassuspension system need not necessarily be used in isolation but mayinstead be used in conjunction with springs and/or rubber components ofthe type previously described.

As has been previously stated, ships typically posess two chains andchain lockers disposed close to the bow and each of their masses may beused in one of two seperate apparatus designed to reduce the amplitudeof two seperate modes of wave-induced vibrations in the ship's hull.However one of the disadvantages with this arrangement is that the massin any one of the two apparatus is effectively halved while in additionthe independant oscillation of two such masses at the bow of the shipmay cause a twisting motion detrimental to the ship's performance.

It will be apparant to those skilled in the art that whilst anembodiment of the present invention has been described with particularreference to the two node mode of wave-induced vibrations, the apparatusof the present invention is capable of damping any desired mode of hullvibration.

What is claimed:
 1. An apparatus for inclusion in a ship comprising amass at least in part comprised of a chain locker and that portion ofthe chain stored therein and having a hull vibrating at fundamental andharmonic frequencies of a transverse two node mode of wave inducedvibration, means for vibrationally coupling the mass to the ship's hulland means for damping the movement of the mass, the mass being so sizedand located as to, when in use, reduce the fundamental and/or harmonicfrequencies of the transverse two node mode of wave-induced vibration inthe ship's hull.
 2. An apparatus in accordance with claim 1, wherein themass is in part comprised of tackle, fitments or cargo of the ship. 3.An apparatus for inclusion in a ship having a hull, the apparatuscomprising a mass at least in part comprised of a chain locker and thatportion of the chain stored therein, means for vibrationally couplingthe mass to the ship's hull and means for damping the movement of themass so as to, when in use, thereby reduce vibrations in the ship'shull.
 4. An apparatus in accordance with claim 3, wherein the masscomprises between approximately 0.2% and approximately 0.5% of theunloaded tonnage of the ship.
 5. An apparatus in accordance with claim3, wherein the means for vibrationally coupling the mass to the ship'shull comprises a resilient support.
 6. An apparatus in accordance withclaim 3, wherein the means for vibrationally coupling the mass to theship's hull comprises at least one spring.
 7. An apparatus in accordancewith claim 3, wherein the means for vibrationally coupling the mass tothe ship's hull comprises a hydragas suspension system.
 8. An apparatusin accordance with claim 7, wherein the hydragas suspension systemcomprises at least two cylinders disposed in opposed relationship.
 9. Anapparatus in accordance with claim 3, wherein the mass is coupled to theship's hull in such a way as to vibrate at a frequency within 10% of theresonant frequency of a mode of vibration that is to be damped.
 10. Anapparatus in accordance with claim 3, wherein the mass is coupled to theship's hull in such a way as to vibrate when the ship is unloaded at afrequency less than the resonant frequency of a mode of vibration thatis to be damped.
 11. An apparatus in accordance with claim 3, whereinthe frequency of vibration of the mass may be varied by altering thestiffness with which the mass is vibrationally coupled to the ship'shull.
 12. An apparatus in accordance with claim 3, wherein the frequencyof vibration of the mass may be varied by altering the magnitude of themass.
 13. An apparatus in accordance with claim 3, wherein the means fordamping the movement of the mass comprises at least one shock absorber.14. An apparatus in accordance with claim 3, wherein the means fordamping the movement of the mass comprises at least one piston damper.15. A ship having a hull which vibrates in a transverse two node modefrom wave-induced vibration having a fundamental frequency and harmonicfrequencies incorporating an apparatus comprising a mass at least inpart comprised of a chain locker and that portion of the chain storedtherein, means for vibrationally coupling the mass to the ship's hulland means for damping the movement of the mass, the mass being so sizedand located as to, when in use, reduce at least one of the fundamentaland harmonic frequencies of the transverse two node mode of wave-inducedvibration in the ship's hull.
 16. A ship in accordance with claim 15,wherein the mass is in part comprised of at least one of a plurality ofcomponents comprising tackle, fitments and cargo of the ship.
 17. A shiphaving a hull and incorporating an apparatus comprising a mass at leastin part comprised of a chain locker and that portion of the chain storedtherein, means for vibrationally coupling the mass to the ship's hulland means for damping the movement of the mass so as to thereby reducevibrations in the ship's hull.
 18. A ship in accordance with claim 17,wherein the apparatus is located at a point at which the amplitude of amode of vibration that is to be damped is significant.
 19. A ship inaccordance with claim 17, wherein the apparatus is located at a pointwithin one eighth of a wavelength of an antinode of a mode of vibrationthat is to be damped.
 20. A ship in accordance with claim 17, whereintile apparatus is located within a bow region of the ship.
 21. A methodof reducing at least one of fundamental and harmonic frequencies of atransverse two node mode of wave-induced vibration in a ship's hullcomprising the steps of providing a mass of a sufficient size whereinthe mass is at least in part comprised of a chain locker and that partof the chain stored therein, vibrationally coupling the mass to theship's hull at an appropriate location and damping the resultingmovement of the mass.
 22. A method of reducing vibrations in a ship'shull comprising the steps of identifying a mass at least in partcomprised of a chain locker and that part of the chain stored therein,vibrationally coupling the mass to the ship's hull and damping theresulting movement of the mass.
 23. A method in accordance with claim21, wherein the mass is coupled to the ship's hull in such a way as tovibrate at a frequency within 10% of the resonant frequency of a mode ofvibration that is to be damped.
 24. A method in accordance with claim21, wherein tile mass is coupled to the ship's hull in such a way as tovibrate when the ship is unloaded at a frequency less than the resonantfrequency of a mode of vibration that is to be damped.
 25. A method inaccordance with claim 21, wherein the frequency of vibration of the massmay be varied by altering the magnitude of the mass.
 26. A method inaccordance with claim 21 wherein the frequency of vibration of the massmay be varied by altering the stiffness with which the mass isvibrationally coupled to the ship's hull.